Refrigerant flow management center for automobiles

ABSTRACT

A flow management device is provided for managing refrigerant flow in a reversible HVAC system. Refrigerant lines from the system heat exchangers connect to bi-directional ports on the flow management device which converts the high pressure refrigerant flowing in one bi-directional port into pressure reduced refrigerant that flows out of the other bi-directional port. The flow management device has multiple bi-directional ports and a flow path for refrigerant extending between the ports. Flow sensitive valves are positioned within the flow path to prevent high pressure refrigerant from flowing between the ports. A pressure reducing device is arranged so that high pressure refrigerant flowing into one of the ports flows through one of the flow sensitive valves and then into the pressure reducing device. Pressure reduced refrigerant emitted from the pressure reducing device flows through a multi-function valve and out the other port. The invention can further integrate the receiver/drier function into the flow management device providing a centralized device for filtering refrigerant flow and ensuring a continuous supply of liquid refrigerant to the pressure reducing device. Also, outlets can be added to the receiver portion of the flow management center to provide a source of high pressure liquid refrigerant for secondary heat exchangers.

BACKGROUND AND SUMMARY OF THE INVENTION

The present invention relates generally to automotive HVAC systems forcontrolling the environment of an automobile passenger compartment. Moreparticularly, the invention relates to a flow management system forcontrolling refrigerant flow in an automotive HVAC system.

This application is related to co-pending applications all filed on Nov.12, 1998 and titled Reversible Air Conditioning And Heat Pump HVACSystem For Electric Vehicles, Controller For Reversible Air ConditioningAnd Heat Pump HVAC System For Electric Vehicles, Anti-Fog Controller ForReversible Air Conditioning And Heat Pump HVAC System For ElectricVehicles, Controller For Heating In Reversible Air Conditioning And HeatPump HVAC System For Electric Vehicles, Air Handling Controller For HvacSystem For Electric Vehicles, and System For Cooling Electric VehicleBatteries. Each of these applications is incorporated by reference intothe present application.

Automotive climate control systems have traditionally been single loopdesigns in which the full volume of refrigerant flows through eachcomponent in the system. In operation, refrigerant in the vapor phase ispressurized by a compressor or pump. The pressurized refrigerant flowsthrough a condenser which is typically configured as a long serpentinecoil. As refrigerant flows through the condenser, heat energy stored inthe refrigerant is radiated to the external environment resulting in therefrigerant transitioning to a liquid phase. The liquefied refrigerantflows from the condenser to an expansion valve located prior to anevaporator. As the liquid flows through the expansion valve it isconverted from a high pressure, high temperature liquid to a lowpressure, low temperature spray allowing it to absorb heat. Therefrigerant flows through the evaporator absorbing heat from the airthat is blown through the evaporator fins. When a sufficient amount ofheat is absorbed the refrigerant transitions to the vapor phase. Anyfurther heat that is absorbed raises the vaporized refrigerant into asuperheated temperature range where the temperature of the refrigerantincreases beyond the saturation temperature. The superheated refrigerantflows from the outlet of the evaporator to the compressor where thecycle repeats.

Generally, the refrigerant flowing into the compressor should be in thevapor phase to maximize pumping efficiency. The operation of therefrigerant loop in conventional automotive HVAC systems is controlledby cycling the compressor on and off as well as by varying the volume ofrefrigerant that is permitted to flow through the expansion valve.Increasing the volume of refrigerant that flows through the valvelengthens the distance traversed by the liquid before it changes to thevapor phase, allowing the heat exchanger to operate at maximumefficiency.

Advances in automotive HVAC systems have led to zone temperature controlsystems in which different zones of an automobile are independentlycontrolled. Zone control systems generally include an evaporator andexpansion valve for each zone. The refrigerant flows through acompressor and condenser, then is split by a system of valves beforeflowing to the expansion valve and evaporator of each zone. Therefrigerant flowing out of the evaporator of each zone is thenrecombined before returning to the compressor. A complex series ofvalves and plumbing is generally required to maintain a balanced HVACsystem that provides individualized control for each of the zones. Therefrigerant plumbing associated with zone control systems issignificantly more complex than the plumbing of prior single loopdesigns.

The complexity of refrigerant plumbing has further increased with therecent implementation of reversible heat pump systems in automobiles. Ina reversible heat pump system the direction of the refrigerant flow iscontrolled by a four-way valve, thus permitting the HVAC system tooperate in both a heating mode and a cooling mode. In the cooling moderefrigerant flows from the compressor through an outside coil(condenser) and into an expansion valve and inside coil (evaporator)before returning to the compressor. Heat energy is extracted from airthat is blown through the inside coil (evaporator), thus providingcooled air to the passenger compartment. In the heating mode the fourway switch reverses the flow of refrigerant through the coils, therebyreversing the function of the coils. Refrigerant flows from thecompressor through the inside coil (condenser) then into an expansionvalve and the outside coil (evaporator) before returning to thecompressor. Heat energy in the liquefied refrigerant flowing through theinside coil is absorbed by air that is blown through the coil into thepassenger compartment thus providing heated air. For the reversiblesystem to operate, valves with associated plumbing must be provided tobypass one of the expansion valves during each mode. When zone controlis added to a reversible heat pump system the complexity and cost of theHVAC further increases. In addition to excessive cost, the systembecomes less reliable due to the increased number of valves, plumbing,and control software required for the system.

One object of the present invention is to provide a flow managementdevice with bi-directional ports in which refrigerant flows into one ofthe ports passes through an expansion valve, and exits the other port.

Another object of the present invention is to integrate a receiver/drierfunction into the flow management device to provide a continuous sourceof liquid refrigerant to the pressure reducing device in the flowmanagement center.

A further object of the present invention is to integrate outlets in theflow management device from which refrigerant can be obtained forsecondary cooling circuits.

It is now proposed, according to the present invention, to provide aflow management device to reduce the cost and improve the performanceand reliability of automotive reversible HVAC systems by simplifying theinterconnection of the HVAC components. Refrigerant lines from thesystem heat exchangers connect to bi-directional ports on the flowmanagement device which converts the high pressure refrigerant flowingin one bi-directional port into pressure reduced refrigerant that flowsout of the other bi-directional port. The flow management device hasmultiple bi-directional ports and a flow path for refrigerant extendingbetween the ports. Flow sensitive valves are positioned within the flowpath to prevent high pressure refrigerant from flowing between theports. A pressure reducing device is arranged so that high pressurerefrigerant flowing into one of the ports flows through one of the flowsensitive valves and then into the pressure reducing device. Pressurereduced refrigerant emitted from the pressure reducing device flowsthrough a multi-function valve and out the other port. The inventionsimplifies the interconnection of reversible HVAC systems by eliminatingthe complex plumbing and extra valves associated with conventionalsystems. In addition, the invention can further integrate thereceiver/drier function into the flow management device providing acentralized device for filtering refrigerant flow and ensuring acontinuous supply of liquid refrigerant to the pressure reducing device.Also, outlets can be added to the receiver portion of the flowmanagement center to provide a source of high pressure liquidrefrigerant for secondary heat exchangers.

The above described device is only an example. Devices in accordancewith the present invention may be implemented in a variety of ways.

BRIEF DESCRIPTION OF THE DRAWINGS

These and other objects of the present invention will become apparent tothose skilled in the art from the following detailed description inconjunction with the attached drawings in which:

FIG. 1 is a schematic representation of a preferred embodiment of theautomotive refrigerant circuit;

FIG. 2 is a cross-sectional view of the flow management center shown inFIG. 1;

FIGS. 3a and 3b present cross-sectional views of flow management devicesembodying the present invention;

FIG. 4 is a schematic representation of an alternative automotiverefrigerant circuit;

FIG. 5 is a block diagram illustration of the control circuitinterconnection to a reversible HVAC refrigerant circuit;

FIG. 6 is a flow diagram showing an overview of the control program forthe preferred embodiment of the invention;

FIG. 7 is a flow diagram illustration of the expansion valve controlprogram for the preferred embodiment of the invention;

FIG. 8 is a flow diagram of the compressor speed control module for thepreferred embodiment of the invention;

FIG. 9 is a diagram illustrating the interaction between the expansionvalve and compressor during the turn-on transition;

FIG. 10 is a datagram illustrating the relationship between thetemperature cycle and a schematic representation of an HVAC system;

FIG. 11 is a flow diagram of the anti-fog algorithm for the preferredembodiment of the invention;

FIG. 12 is a flow diagram of the heating mode selection module for thepreferred embodiment of the invention;

FIG. 13 is a flow diagram of the air-handling method for the preferredembodiment of the invention;

FIG. 14 is a schematic representation of a preferred embodiment of anHVAC system coupled to a battery pack module;

FIG. 15 is datagram illustrating the relationship between a preferredembodiment of the HVAC system and its heat load cycle; and

FIG. 16 is a datagram illustrating the relationship between thetemperature lever position and the corresponding operating mode.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

FIG. 1 illustrates an exemplary reversible HVAC system 50 for motorvehicles that includes an air-flow structure 52, a refrigerant flowsystem 54, and a front panel 55 for providing controlling inputs. Thereversible HVAC system 50 can both heat and cool the passengercompartment air of a motor vehicle by using the refrigerant flow system54 in conjunction with the air-flow structure 52 to transfer heat energybetween the outside environment and the passenger compartment. Inheating mode, heat energy is transferred from the outside environment toair that flows into the passenger compartment and in cooling mode, heatenergy is transferred to the outside environment from air that flowsinto the passenger compartment. The refrigerant flow system 54 acts as astorage medium for heat energy that is being transferred between theoutside environment and the passenger compartment. The air-flowstructure 52 controls the flow of conditioned air into the passengercompartment. An inside heat exchanger 88 provides an interface betweenthe refrigerant flow system 54 and the air-flow structure permitting thetransfer of heat energy between the refrigerant and the air flowing intothe passenger compartment. The front panel 55 provides a means for thepassengers to control the temperature, flow rate, and operating mode ofthe HVAC system.

The air-flow structure 52 includes a duct 56 through which air issupplied into the passenger compartment, a blower 58 for introducing airinto the duct 56, a recirculation door 60 for controlling the proportionof fresh air to recirculated air, a PTC heater 62 for heating the air, ablend door 60 for controlling the proportion of air that flows over thePTC heater 62, and a set of duct outlets for discharging air into thepassenger compartment.

The duct outlets include a defrost outlet 64 for directing air towardsthe windshield of the vehicle, a panel outlet 66 for directing airtowards the upper extremities of the passengers, and a floor outlet 68for discharging air towards the lower extremities of the passengers. Theduct outlets 64-68 are selectively opened and closed by a mode damper 70which operates in accordance with the position of the mode selectorswitch 72 located on front panel 55.

The refrigerant flow system 54 is operable in a heating mode and acooling mode and includes a compressor 76, a four-way switch 78 forcontrolling the direction of refrigerant flow, an inside heat exchanger88 for transferring energy between the refrigerant and air flowing intothe passenger compartment, an outside heat exchanger 80 for interfacingwith the outside environment, a flow management center 82 for reducingthe pressure of refrigerant flowing into a heat exchanger that isfunctioning as an evaporator, shut-off valves 84 and 86 for systemprotection, zone-control heat exchanger 92 for providing independentlycontrolled cooling to a local region, and pressure reducing device 90for reducing the pressure of refrigerant flowing into the zone-controlheat exchanger 92. The refrigerant flow system 54 interacts with theair-flow structure 52 and the passenger compartment through theoperation of the inside heat exchanger 88 during the heating and coolingmodes. The function of the inside heat exchanger 88 changes in eachoperating mode; during heating mode the inside heat exchanger 88functions as a condenser transferring heat energy to air that passesthrough air-flow structure 52 into the passenger compartment and duringcooling mode the inside heat exchanger 88 functions as an evaporatorabsorbing heat energy from the air that passes through air-flowstructure 52 into the passenger compartment.

The compressor 76 is driven by a variable speed electric motor (notshown). Varying the speed of the electric motor causes a commensuratechange in the suction pressure and refrigerant discharge capacity ofcompressor 76. Although the compressor in the present embodiment is avariable speed compressor, it is within the scope of the invention toemploy a single speed compressor. The four-way switch 78 is connectedbetween the compressor 76 and the heat exchangers 80 and 88 to provide amethod of changing from air conditioning mode to heat pump mode byreversing the direction of refrigerant flow.

The inside heat exchanger 88 functions as an evaporator during a coolingoperation and as a condenser during a heating operation. Inside heatexchanger 88 is arranged within duct 56 so that the air blown throughthe exchanger 88 is conditioned prior to passing over PTC heater 62 andbeing discharged through the duct outlets. Shut-off valve 84 provides ameans of interrupting refrigerant flow during HVAC operating modes thatdo not require operation of inside heat exchanger 88. Examples of suchoperating modes include disabling operation of the inside heat exchanger88 as an evaporator at low ambient temperatures that could result infreezing of the heat exchanger 88 due to condensation, and modes whereonly secondary heat exchangers are operational such as zone control heatexchanger 92. Such operating modes include cooling of a battery assemblyand cooling of pre-selected regions within the vehicle. The flowmanagement center 82 reduces the pressure of and expands the refrigerantto be supplied to the inside heat exchanger 88 during a coolingoperation.

The outside heat exchanger 80, which is generally located towards thefront of the vehicle, exchanges heat between the outside air and therefrigerant. A fan 94 ensures a constant supply of outside air flowsthrough outside heat exchanger 80. During air conditioning mode theoutside heat exchanger 80 functions as a condenser providing a means forthe refrigerant to shed heat to the outside air. During heat pump modethe outside heat exchanger 80 functions as an evaporator absorbing heatenergy from the outside air into the refrigerant.

The flow management center 82 provides a centrallized device forreducing the pressure of refrigerant flowing into a heat exchanger 80 or88 functioning as an evaporator and acts as a source of high pressureliquid refrigerant for secondary heat exchangers. Conventional circuitsuse a separate pressure reducing device with bypass plumbing for eachheat exchanger that functions as an evaporator. By using a single flowmanagement center 82 to provide pressure reduced refrigerant thecomplexity of the refrigerant flow system 54 is greatly reduced.Additionally, a receiver/drier function is integrated into the flowmanagement center 82 for eliminating contaminants and providing areservoir of pressurized liquid refrigerant. Refrigerant tapped from thereceiver portion is routed to pressure reducing device 90 and then tozone-control heat exchanger 92. Although the flow management center inthe preferred embodiment includes a receiver/drier function theprinciples of the invention can be extended to flow management devicesthat do not include a receiver/drier function.

Flow management center 82 is illustrated in greater detail in FIG. 2 toinclude a housing 100 defining bi-directional ports 102 and 104, apressure sensitive valve 106, check valves 108 and 110, desiccant 112, auni-directional flow member 114, pressure reducing valve 116, outlets118 and 120, temperature probe 124, and pressure probe 122. Pressurizedliquid refrigerant flows into bi-directional port 102 or 104, throughthe corresponding check valve 108 or 110, through the dessicant 112,into reservoir 113, up the uni-directional flow member 114, throughpressure reducing device 116 and pressure sensitive valve 106, andfinally reduced pressure refrigerant flows out of the otherbi-directional port 104 or 102. When the HVAC system 50 changesoperating modes the direction of refrigerant flow reverses as highpressure refrigerant flows into the bi-directional port that pressurereduced refrigerant was flowing from. The refrigerant then flows throughthe corresponding check valve 110 or 108, through the dessicant 112,into reservoir 113, up the uni-directional flow member 114, throughpressure reducing device 116 and pressure sensitive valve 106, andfinally reduced pressure refrigerant flows out of the otherbi-directional port 102 or 104. The pressure sensitive valve 106 permitsthe flow of pressure reduced refrigerant out of one bi-directional portwhile preventing high pressure refrigerant from flowing directly betweenthe bi-directional ports. When high pressure refrigerant flows into abi-directional port 102 and 104 the pressure sensitive valve 106 closesthe flow path from the port to the pressure reducing device 116 andopens a path from the pressure reducing device to the otherbi-directional port 104 and 102. Closing the flow path from thebi-directional port 102 or 104 to the pressure reducing device forcesrefrigerant to flow through the corresponding check valve 108 or 110,through the dessicant 112, and into reservoir 113. The opposing checkvalve 110 or 108 prevents high pressure liquid refrigerant in reservoir113 from flowing out the opposing bi-directional port 104 or 102.Impurities within the refrigerant are removed by dessicant 112.Reservoir 113 provides a pool of high pressure liquid refrigerant thatcan be sourced to multiple pressure reducing devices such as device 116within the flow management center 82 as well as pressure reducingdevices that provide reduced pressure refrigerant to secondary heatexchangers. Outlets 118 and 120 provide a means of tapping offrefrigerant from reservoir 113 and directing it to secondary heatexchanger circuits. In the preferred embodiment the pressure sensitivevalve 106 is a dual poppet valve, however it is envisioned that othervalves such as multiple check valves, mushroom valves, reed valves, orrotary valves may be employed. Additionally, similar valves as listedabove can replace check valves 108 and 110. Although the pressurereducing device 116 in the preferred embodiment is an electronicallycontrolled expansion valve it is within the scope of the invention touse mechanically controlled expansion valves as well as 90° valves. Thedesiccant 112 and the temperature and pressure probes 122 and 124 aremerely exemplary of additional functions that can be added to the flowmanagement center, they are not required to practice the invention.

Returning to FIG. 1, the zone-control heat exchanger 92, located withinthe interior of the vehicle provides cooling functions for local zonesor assemblies. Examples of local zone cooling include battery assemblycooling, air conditioned seats, and individualized cooling of one sideof the passenger compartment. Pressure reducing device 90 reduces thepressure of and expands the refrigerant to be supplied to zone controlheat exchanger 92. The expanded refrigerant absorbs heat from the air orliquid which is passed through heat exchanger 92, thereby cooling theair or liquid.

The front panel 55 includes selector switches for setting the operatingparameters of the air conditioning circuit 50. The switches include ablower speed selector 73 that in the preferred embodiment is adjustablefrom 30% to 100% of the maximum blower speed, a mode selector switch 72having five mode settings, a recirculation selector 75 for selectingfresh or recirculated air, and a sliding temperature lever 74 forsetting the temperature of air discharged from the duct outlets.Although the mode selector switch in the preferred embodiment has fivediscrete settings, the principles of the invention can be extended to amode selector having an unlimited number of settings.

During cooling mode, the refrigerant discharged from the compressor 76flows through four-way switch 78 into outside heat exchanger 80 whichfunctions as a condenser. As heat energy stored in the refrigerant isshed to the outside air which is blown through the exchanger 80 therefrigerant condenses to a high pressure liquid. The liquid refrigerantflows into a bi-directional port 102 of the flow management center 82,through the desiccant 112, into the reservoir 113, up theuni-directional flow member 114, through the pressure reducing valve116, and then out the other bi-directional port 104. A portion of therefrigerant is tapped off from the reservoir 113 and directed towards asecondary loop as shown in FIG. 1 will be explained in a laterparagraph. The refrigerant flowing through the pressure reducing valve116 is pressure reduced and then passes through the other bi-directionalport 104. The pressure reduced refrigerant flows into the inside heatexchanger 88 which functions as an evaporator. Heat energy from airpassing through inside heat exchanger 88 is absorbed by the pressurereduced refrigerant causing the refrigerant to change to the vaporstate. The vapor state refrigerant flows from the heat exchanger 88through the four-way switch 78 and back to the inlet of compressor 76which compresses the vapor and directs it through four-way switch 78 tooutside heat exchanger 80.

The operation of the secondary loop during cooling mode is as follows.The portion of refrigerant that flowed from an outlet in reservoir 113flows through shut-off valve 86 into pressure reducing device 90.Pressure reduced refrigerant flows out of device 90 into local-zone heatexchanger 92 which functions as an evaporator. The refrigerant absorbsheat from the air which passes through it thereby providing separatelycontrolled cooling for a portion of the passenger compartment. Althoughthe zone control heat exchanger 92 in the preferred embodiment functionsas an air-to-refrigerant evaporator, it is within the scope of theinvention to employ other heat exchangers such asrefrigerant-to-refrigerant, water-to-refrigerant, and oil-to-refrigerantheat exchangers.

During heating mode, the direction of refrigerant flow is reversed bychanging the orientation of four-way switch 78. A signal from acontroller 130, hereinafter described, controls the orientation offour-way switch 78. The refrigerant discharged from the compressor 76flows through four-way switch 78 into inside heat exchanger 88 whichfunctions as a condenser. As heat energy stored in the refrigerant isshed to the inside air which is blown through the exchanger 88 therefrigerant condenses to a high pressure liquid. The liquid refrigerantflows into the bi-directional port 104 of the flow management center 82,through the desiccant 112, into the reservoir 113, up theuni-directional flow member 114, through the pressure reducing valve116, and then out the other bi-directional port 102. The refrigerantflowing through the pressure reducing valve 116 is pressure reduced andthen passes through bi-directional port 102. The pressure reducedrefrigerant flows into the outside heat exchanger 80 which functions asan evaporator. Heat energy from air passing through outside heatexchanger 80 is absorbed by the pressure reduced refrigerant causing therefrigerant to change to the vapor state. The vapor state refrigerantflows from the heat exchanger 80 through the four-way switch 78 and backto the inlet of compressor 76 which compresses the vapor and directs itback through four-way switch 78 to inside heat exchanger 88.

During heating mode, the secondary loop operates in the same manner asduring a cooling mode. Refrigerant from outlet 118 of flow managementcenter 82 flows through pressure reducing device 90 and into local-zoneheat exchanger 92 in which it absorbs heat from air that is passingthrough the exchanger 92. Pressure reducing device 90 pressure reducesthe refrigerant to increase its capacity to absorb heat energy from airor fluid flowing through the heat exchanger 92.

Employing flow management center 82 in reversible HVAC system 50 greatlysimplifies the interconnecting plumbing and permits more reliableimplementation of secondary cooling loops. It is possible to alternatelyheat and cool a vehicle with two heat exchangers without the additionalvalves and plumbing required for conventional systems. Complexrefrigerant balancing schemes for dividing refrigerant amongst multipleheat exchanger loops are not required, thereby improving systemperformance, increasing system reliability, and reducing cost. A commonsense point at the outlet of pressure reducing device 116 is providedfor pressure reduced (low-side) refrigerant. Sensing temperature andpressure at the flow management center eliminates the need ofconventional systems for sensing at the inlet to each heat exchanger.

Referring to FIGS. 3a and 3b, an alternate flow management device 81 isillustrated which does not include the receiver/drier function, butprovides reversibility with simpler plumbing than conventional systemsand a single pressure reducing device. The flow management deviceincludes a housing 100 defining bi-directional ports 102 and 104, apressure sensitive valve 106, check valves 108 and 110, auni-directional flow member 114, temperature probe 122, and pressureprobe 124. The flow management device 81 includes all the capabilitiesof the flow management center 82 with the exception of thereceiver/drier function.

FIG. 4 illustrates another embodiment of an automotive air conditioningcircuit 40 that includes a compressor 41, an outside heat exchanger 42,an inside heat exchanger 43, two four-way switches 44 and 45, areceiver/drier 46, and an electronic expansion valve 47.

Four-way switch 45, receiver/drier 46, and expansion valve 47functionally replace the flow management center 82 that is employed incircuit 50 (see FIG. 1). The function of four-way valve 45 is the mirrorimage of the function of four-way valve 44. Valve 44 is employed toreverse the flow of refrigerant through the heat exchangers 42 and 43.It essentially converts uni-directional refrigerant flow from thecompressor 41 into bi-directional refrigerant flow into the heatexchangers 42 and 43. Whereas four-way valve 45 converts bi-directionalrefrigerant flow from the heat exchangers 42 and 43 into auni-directional flow through receiver/drier 46 and expansion valve 47.

Receiver/drier 46 removes contaminants from the refrigerant and ensuresa continuous flow of high pressure liquid refrigerant into expansionvalve 47. Expansion valve 47 provides refrigerant pressure reduction andexpansion for heat exchangers 42 and 43. Expansion valve 47 ispreferably an electronic expansion valve that receives its controllinginputs from a controller that monitors the saturation and superheattemperature of the heat exchangers 42 and 43. However, other pressurereducing devices such as block valves, 90° valves, and thermal expansionvalves (TXV) are within the scope of the invention. Generally, tocontrol a TXV, refrigerant at the superheat temperature and thesaturation temperature must be routed to the device. To obtain thesuperheat temperature the refrigerant from four-way valve 44 to thecompressor 41 inlet can be routed through the TXV. For the saturationtemperature the refrigerant emitted from the TXV can be sensed.

During cooling mode outside heat exchanger 42 functions as a condensershedding heat to the outside environment and inside heat exchanger 43functions as an evaporator absorbing heat from air that is blown intothe passenger compartment. The refrigerant cycle is as follows:refrigerant flows out of compressor 41, through four-way valve 44, intothe outside heat exchanger 42, through four-way valve 45, intoreceiver/drier 46 and expansion valve 47, through four-way valve 45, toinside heat exchanger 43, through four-way valve 44, and back tocompressor 41.

During heating mode four-way valve 44 changes orientation causing theflow of refrigerant to heat exchangers 42 and 43 to reverse. With thereversal in the direction of refrigerant flow the functions of the heatexchangers 42 and 43 reverse as inside heat exchanger 43 functions as acondenser and outside heat exchanger 42 functions as an evaporator. Inaddition, the orientation of four-way valve 45 is also changed to ensurethat the direction of refrigerant flowing into receiver/drier 46 andexpansion valve 47 remains constant. The refrigerant cycle during heatpump mode is as follows: refrigerant flows out of compressor 41, throughfour-way valve 44, into the inside heat exchanger 43, through four-wayvalve 45, into receiver/drier 46 and expansion valve 47, throughfour-way valve 45, to outside heat exchanger 42, through four-way valve44, and back to compressor 41.

From the foregoing it will be understood that the invention provides aflow management device with bi-directional ports in which refrigerantflowing into either port passes through an expansion valve and exits theother port. Additionally, the invention can integrate the receiver/drierfunction into a flow management device with bi-directional ports toprovide the capability of tapping off refrigerant flow for secondarycooling circuits. Also, the present invention decreases the complexityof automotive HVAC systems by integrating a flow management device intothe system to reduce the number of valves required to implement areversible heating and cooling HVAC system. A further capability of theinvention is to provide a centralized flow management center with tapsfor refrigerant to reduce the complexity of automotive HVAC systems thatimplement multi-zone control.

Control System for Reversible Air Conditioning and Heat Pump HVAC Systemfor Electric Vehicles

FIG. 5 illustrates the control system configuration to implement thepreferred embodiment of the HVAC circuit 50. In FIG. 5 the outside coil80, flow management center 82, inside heat exchanger 88, four-way switch78, compressor 76, duct 56, and front panel 55 are interconnected in amanner similar to circuit 50 illustrated in FIG. 1. Additionallyillustrated is controller 130 which controls the compressor speed andflow management center 82 operation based upon inputs from front panel55, duct 56, and the refrigerant system 54.

During operation of the HVAC circuit 50, the passenger selects apassenger compartment temperature and operating mode by setting theswitches of front panel 55. The front panel 55 switch settings aredecoded by the controller 130, which converts the settings to valuesthat represent desired temperature, operating mode, and blower speed.The controller 130 also monitors sensors that measure the actual ambientand passenger compartment temperature as well as refrigerant temperatureand pressure. The controller 130 compares the decoded settings to theactual ambient and passenger compartment temperature, and generatessignals that modify the operation of the refrigerant flow system 54 andair-flow structure 52 to bring the actual passenger compartmenttemperature in conformance with the desired temperature as representedby the front panel 55 switch settings.

The operation of the refrigerant flow system 54 is modified bycontroller 130 through output signals that control the orientation ofthe four-way switch 78, the speed of compressor 76 and the duty cycleapplied to the pressure reducing device 116 within the flow managementcenter 82. Changing the orientation of four-way switch 78 causes areversal in the direction of refrigerant flow. The direction thatrefrigerant flows dictates whether the HVAC system is in the heatingmode or the cooling mode by interchanging the functions of the outsideheat exchanger 80 and the inside heat exchanger 88. In heating mode theoutside heat exchanger 80 functions as an evaporator and the inside heatexchanger 88 functions as a condenser 88. Whereas, in cooling mode theoutside heat exchanger 80 functions as a condenser and the inside heatexchanger 88 functions as an evaporator. Varying the speed of compressor76 during a cooling mode or a heating mode causes a change in therefrigerant temperature at the compressor 76 inlet and outlet, which hasa direct effect on the temperature of air blown into the passengercompartment. Changing the duty cycle applied to the pressure reducingdevice 116 during either cooling or heating mode causes a variation inthe quantity of refrigerant that the pressure reducing device 116permits to flow into the heat exchanger 80 or 88 that is functioning asan evaporator. Too much refrigerant flowing through the evaporator leadsto flooding the compressor 76, causing degraded compressor 76performance. Too little refrigerant flowing through the evaporatorlimits the efficiency of the evaporator in absorbing heat, resulting ina reduced cooling or heating capacity of the HVAC system 50. Thecontroller 130 constantly adjusts the duty cycle applied to the pressurereducing device to keep the evaporator operating at maximum efficiencyand adjusts the speed of compressor 76 to control the temperature of theair blown into the passenger compartment.

The air-flow structure 52 operation is modified by changing the positionof blend door 61 and the position of recirculation door 60. Changing theposition of blend door 61 changes the amount of supplemental electricheating that is applied to the air flowing through the air-flowstructure 52, directly effecting the temperature of the passengercompartment. The position of recirculation door 60 controls whetherfresh air from the outside or recirculated air from inside is directedinto the passenger compartment. Typically, more energy is required toheat or cool fresh air than recirculated air because of the greaterdifferential between the temperature of the air flowing into the HVACsystem 50 and the desired passenger compartment temperature.

Inputs to controller 130 from the front panel 55 include blower speedfrom blower speed selector 73, mode selection from mode selector switch72, and the target temperature from temperature lever 74. The duct 56inputs include inlet and outlet temperatures from temperature probes132, 133, and 134. Inputs from the refrigerant system 54 to thecontroller 130 include temperature probe 135 for sensing ambienttemperature, temperature probe 124 for sensing the expansion valve 116outlet temperature, temperature probe 136 for sensing superheattemperature, and pressure probe 138 for sensing suction pressure.

Controller 130 is preferably a microprocessor-based controller, thatincludes a processor 140 and associated memory 142. An analog-to-digitalconverter (A/D) 144 converts signals from the various sensors to adigital form used by processor 140. A driver circuit 146 operates theflow management center 82 and compressor 76. This may be for example aninterface circuit that connects to the electric motor for driving thecompressor 76 in response to system temperature inputs. The interfacecircuit may also provide a duty cycle signal for controlling theexpansion valve 116 to maintain a regulated average superheattemperature in the compressor suction line. Additionally, the driver 146may include an interface circuit coupled to four-way switch 78 forreversing the switch from cooling mode to heating mode.

Processor 140 includes a main program 151, depicted in the flowchart ofFIG. 6, to control the operating mode selection, compressor speedcontrol, and electronic expansion valve (EXV) control. FIG. 6 gives anoverview of the control strategy illustrating the major functionalmodules that are involved.

Referring to FIG. 6, the main program 151 is illustrated. The mainprogram 151 provides the timing for execution of the various controlmodules. At step 152 the program enters the operating mode selectionmodule in which the operating mode of the system is selected. Thesupported operating modes include defrost mode, vent mode, PTC heatermode, heat pump mode, and air conditioning mode. The inputs monitored bythe controller 130 to select the HVAC system 50 operating mode includethe position of mode selector switch 72, temperature lever 74, inlettemperature, and during a heating operation the capacity of heat pumpmode. Although the preferred embodiment has five discrete operatingmodes, the principles of the invention can be extended to systems havingeither fewer operating modes or a continuously variable set of operatingmodes.

FIG. 16 illustrates the system operating modes. During the PTCheater/defrost mode, when the ambient temperature is less than 40° F.,controller 130 turns on the PTC heater 62 and moves the blend door to aposition determined by the location of temperature lever 74. However,for the first 3% of temperature lever 74 travel from the full coldposition the controller turns off PTC heater 62 and only enables thevents.

In the heating mode, with ambient temperatures greater than 40° F. ordefrost operation with ambient temperatures between 40° F. and 60° F.,controller 130 turns on the heat pump and if necessary the PTC heaterwith blend door to generate the desired temperature that is reflected bythe position of temperature lever 74. For the first 3% of temperaturelever 74 travel from full cold the controller 130 turns off the heatpump and PTC heater 62 and only enables the vents. At temperaturesgreater than 100° F. the controller 130 turns off PTC heater 62.

The third operating mode, cooling mode, is selectable for ambienttemperatures that are greater than 40° F. Cooling mode is also used fordefrost when the ambient temperature is greater than 60° F. For thefirst 33% of temperature lever travel the controller 130 varies thecompressor suction pressure set point from 20 to 45 psig as thetemperature lever 74 is moved from cold to warm. Varying the suctionpressure set point causes a direct change in the compressor speed,thereby causing the air temperature at the duct outlets to change. From33% to 100% of temperature lever travel the controller 130 sets thecompressor 76 suction pressure to a constant 30 psig and turns on thePTC heater 62 to reheat the conditioned air.

Returning to FIG. 6, at step 154 the program enters the recirculationdoor positioning module which is described below with reference to FIG.13. The recirculation door positioning module controls the proportion offresh air to recirculated air that is blown into the passengercompartment. At steps 156 and 158 the program enters modules formonitoring and disabling the compressor in response to detected faults.The compressor speed control module, which is described below withreference to FIG. 8, is entered at step 160. Varying the speed ofcompressor 76 causes a proportional change in the air temperature blownfrom the duct outlets 64-68. Step 162 leads to the EXV control modulewhich is described with reference to FIG. 7. The EXV control module 162modulates the output of the expansion valve 116 in response to changesin the vapor temperature sensed at the compressor 76 and the compressorsuction pressure. Each of the above listed modules will now be furtherexplained.

FIG. 7 illustrates the detailed operation of EXV control module 162. Themodule 162 controls the volume of refrigerant that is pressure reducedby the expansion valve 116 to maintain a relatively constant superheattemperature at the outlet of the evaporator. As low-pressure refrigerantflows from the expansion valve 116 through the evaporator it absorbsheat from the air passing through the evaporator. After absorbingsufficient heat the low-pressure refrigerant transitions to a vaporstate. Any further heat that is absorbed by the vapor raises therefrigerant temperature above the saturation temperature into asuperheated temperature region. To reduce the outlet temperature of therefrigerant the volume of refrigerant flowing into the evaporator isincreased, thereby increasing the heat load capacity of the refrigerant.However, if there is too great a volume the refrigerant will nottransition to the vapor state, resulting in the compressor 76 beingswamped by liquid refrigerant. An insufficient volume of refrigerantflowing into the evaporator results in the refrigerant transitioning tothe vapor state before reaching the outlet of the evaporator. Vaporstate refrigerant has less capacity to store heat energy than liquidstate refrigerant, therefore the portion of the evaporator that containsvapor state refrigerant has less capacity to store heat energy, reducingthe efficiency of the evaporator. It is desirable to control the EXV 116such that the liquid to vapor transition occurs slightly before theoutlet of the evaporator causing the refrigerant to superheat apredetermined amount. This maximizes the efficiency of the evaporator byensuring that virtually the entire coil is used for absorbing heat.

In step 164 the proportional-integral-differential (PID) constants arechosen based upon whether the system is in heating mode or cooling mode.The selection of PID constants is based upon the particular systemcharacteristics and is well known in the art. Following selection of thePID constants the EXV control module proceeds to steps 166 and 168wherein the expansion valve duty cycle is initialized based upon ambienttemperature and operating mode when the system first enters either heatpump mode or air conditioning mode. The graph appended to step 168depicts the selection criteria for the duty cycle. Ambient temperatureis sensed by temperature probe 135 located in front of the outside heatexchanger 80. The initial duty cycle is then set to a value ranging from50% to 100% of the maximum EXV duty cycle depending on the ambienttemperature. After setting the initial expansion valve duty cycle thesystem transitions through a start-up period before settling intosteady-state operation.

During steady-state operation the duty cycle of the EXV is varied inorder to maintain a constant superheat temperature, 4° F. greater thanthe saturation temperature, at the inlet to compressor 76. At step 170the average superheat temperature is calculated by measuring the vaportemperature of refrigerant exiting the evaporator and subtracting thesaturation temperature of the fluid. The saturation temperature isobtained by measuring the compressor inlet suction pressure and usingthe saturation temperature that corresponds to the suction pressure.Although the present embodiment of the invention calculates the averagesuperheat temperature from the vapor temperature and suction pressure,it is within the scope of the invention to use the vapor temperaturewith an evaporator inlet temperature including compensating for theevaporator pressure drop. The outlet of the expansion valve 116 locatedin the flow management center provides a common temperature measurementlocation for evaporator inlet temperature in either heating mode orcooling mode. In conventional systems that use the evaporator inlettemperature to calculate the superheat temperature; temperature probesare required at the inlets to both the inside and outside heatexchangers to provide inlet temperature in both operating modes.

The updated superheat temperature from step 170 is used at step 172 tocalculate a revised setting for the EXV duty cycle. As a final step, atstep 174 the controller 130 limits the value of the EXV duty cycle tobetween 5% and 100% to ensure the device remains within a knownoperating region.

Referring to FIG. 8, the compressor speed control module 160 isillustrated. The compressor speed is controlled by applying a variableduty cycle to the electric motor that drives the compressor 76. The dutycycle is varied in response to a controlling input such as temperaturelever position and compressor suction pressure. Varying the speed ofcompressor 76 causes a proportionate variation in the dischargetemperature and discharge pressure of refrigerant flowing out of thecompressor 76 as well as an inversely proportional change in thecompressor suction pressure and refrigerant suction temperature. Theincreased refrigerant discharge temperature results in an increasedcondenser temperature, increasing the capacity of the HVAC system 50 toprovide heat during heating mode. The decreased refrigerant suctiontemperature results in a decreased evaporator temperature, increasingthe capacity of the HVAC system 50 provide cooling during cooling mode.The speed of the compressor 76 is therefore varied to maintain air blowninto the passenger compartment at a relatively constant temperatureduring both heating mode and cooling mode.

The desired temperature is set by adjusting the temperature lever 74 onfront panel 55. The controller 130 calculates the target suctionpressure corresponding to the temperature lever position (x/L) which isequal to 20+75*(x/L) for a lever travel distance equal to 33% of theavailable distance. Using the suction pressure as the controlledparameter instead of air temperature provides a more stable and fasterresponding system.

Conventional systems that use air temperature as the controlledparameter have problems with surging of the compressor 76 in addition toslow response time. As the sensed outlet air temperature changes due totransient effects including changes in vehicle speed or passing throughintermittent sunlight, the compressor speed is changed in an attempt tokeep the outlet temperature constant. When the compressor speed isconstantly changing the passenger perceives the changes as surging inthe propulsion of the vehicle. In the preferred embodiment, the EXVcontrol loop regulates a constant outlet temperature while thecompressor regulates a constant suction pressure. As the outlet airtemperature changes the heat that is transferred between the refrigerantand the inside heat exchanger varies, causing the refrigerant superheattemperature to change. In response to the change in the superheattemperature the duty cycle of pressure reducing valve 116 is changed bycontroller 130, causing a shift in the flow of refrigerant, resulting ina slight variation of the compressor suction pressure. The controller130 then modifies the speed of compressor 76 to bring it in conformancewith the target suction pressure. However, the required change in thespeed of the compressor 76 is significantly less than the change thatwould be required in an HVAC system that uses compressor speed alone tocompensate for changes in outlet temperature. The minor change incompressor speed is imperceptible to the passengers, leading to enhanceddriving comfort.

In addition to eliminating surging, the response time of HVAC system 50is reduced by using suction pressure as the controlled input. Coolingair at the desired temperature is blown over passengers in significantlyless time than conventional systems that control air temperaturedirectly. As a result, unlike conventional systems, PTC heating of thecooled air is not required to provide fine control over the airtemperature, resulting in more energy efficient vehicle operation.

During heat mode the compressor speed is varied in reaction to changesin the temperature of the air flowing out of the inside heat exchanger88. In heat pump mode, unlike air conditioning mode, suction pressure isnot directly related to the temperature of air flowing out of the insideheat exchanger. Therefore the air temperature sensed by temperatureprobe 133 is used as the controlling input for compressor speed.

In step 176 the controller 130 calculates the error and error derivativeto be used in the PID controller for the controlled input. In airconditioning mode the controlled input is the suction pressure and inheat pump mode the controlled input is the post inside heat exchangerair temperature measured by temperature probe 133. In step 178 the PIDconstants corresponding to the appropriate operating mode are selected.Then in step 180 the PID controller calculates the change in compressorduty cycle based on the PID constants and the calculated error and errorderivative. The revised duty cycle is limited to between 5% and 90% toensure the compressor 76 is operated within specified parameters.

FIG. 9 illustrates the interaction between the EXV control loop and thecompressor speed control loop during the cooling mode start-uptransition. As explained the EXV control loop regulates the volume ofrefrigerant that flows through pressure reducing device 116 maintaininga predetermined refrigerant superheat temperature at the outlet of theevaporator. A secondary effect of the EXV operation is that as the EXVpermits an increased volume of refrigerant to flow, the suction pressureat the inlet to compressor 76 decreases. The operation of the compressor76 has a corresponding interaction with the EXV. When the speed ofcompressor 76 is changed, the resulting change in suction pressure andtemperature at the inlet to compressor 76 causes a change in thesaturation temperature of refrigerant that flows through the evaporator.Increased compressor 76 speed, causes a lower suction pressure, leadingto a lower saturation temperature, resulting in the refrigeranttemperature rising to the predetermined superheat temperature earlier inthe traverse of the evaporator. The EXV loop compensates for the changein superheat temperature by permitting an increased volume ofrefrigerant to flow through the evaporator, thereby causing a highersuction pressure. When the HVAC system 50 first turns on, if thepressure reducing valve 116 is set to an initial duty cycle of 0%, thevolume of refrigerant flowing through the evaporator will lag thecompressor speed throughout the entire start-up time period, delayingthe start-up, resulting in a start-up time period of approximately 2.5minutes.

Assuming an ambient temperature of 40° F., the EXV is set to an initialduty cycle of 50% at step 168 (see FIG. 7). The compressor suctionpressure is set to achieve the target suction pressure corresponding tothe location of temperature lever 74. Initially, the compressor suctionpressure decreases slightly during the first seconds of operation asfluid pours through the EXV, then as the compressor spins up towardssteady-state speed suction pressure begins to increase significantly. Atthe same time the EXV duty cycle increases until the suction pressurehas increased to a point where the EXV begins to track the suctionpressure. During the early stages of start-up it is not unusual for thecompressor to flood until the compressor speed increases a sufficientamount to develop appropriate suction pressure. In the preferredembodiment the compressor is operated on the borderline of floodingduring the start-up transition thereby contributing to a faster systemresponse time. Also, as the EXV duty cycle begins to track the suctionpressure it will overshoot its steady-state value by a slight amount.The underdamped response displayed by the EXV control loop results in afurther reduction in the system response time. In combination theimprovements result in air cooled to the desired temperature blowingover the faces of passengers within approximately 35 seconds of systemstart-up.

From the foregoing it will be understood that the invention provides asystem for improving the steady-state response time of an automotiveHVAC system. Additionally, the invention permits a reduction in thestart-up time of an automotive air conditioning system. Also, theinvention provides a system for controlling an HVAC system that employsa flow management device. The invention further provides a system forcontrolling an HVAC system incorporating a centralized flow managementcenter.

Anti-Fog System for Reversible Air Conditioning and Heat Pump HVACSystem for Automobiles

Referring to FIG. 10, a single loop reversible air conditioning and heatpump system 191 is illustrated with the corresponding temperature cyclediagrams for air conditioning mode 190 and heat pump mode 192. As willbe described, the preferred embodiment of the present invention preventsundesirable fogging by slowly increasing the speed of compressor 76 overa predetermined period of time. Generally, in reversible HVAC systemsfogging may occur during the transition from cooling mode to heatingmode. Prior to describing the solution provided by the presentlypreferred embodiment, a brief description of the refrigeration cycle andhow fogging occurs in a reversible system is provided with reference toFIG. 10.

The refrigeration cycle essentially uses a small amount of energy topower a compressor in order to transfer a greater amount of heat energyfrom one environmental region to another environmental region. It doesthis by using the cooling effect of evaporation to lower the temperatureof the air passing through one heat exchanger (the evaporator) 88 andusing the heating effect of condensing high temperature, high pressuregas to raise the temperature of the air passing through another heatexchanger (the condenser) 80. With reference to waveform t₁ of FIG. 10,drawn from right to left, the temperature profile of refrigerant flowingfrom an evaporator 88, through a compressor 76 and four-way switch 78,and then through a condenser 80 is illustrated. Refrigerant entering theevaporator 88 is at low pressure and low temperature. The temperaturebeing the saturation temperature of the pressure reduced refrigerant. Asthe refrigerant passes through the evaporator 88 heat energy from airthat is blown through the evaporator 88 is absorbed by the refrigerant.The air that exits the evaporator 88 is noticeably cooled due to thetransfer of heat energy from the air to the refrigerant. The cooler airno longer has the capacity to retain the same amount of moisture as thewarmer air that was blown into the evaporator 88, therefore the excessmoisture condenses out of the air onto the external surface of theevaporator 88. The vapor state refrigerant flows from the evaporator 88to the compressor 76 where it is compressed to a high pressure, hightemperature vapor before flowing into the condenser 80.

When the controller 130 commands a change to heating mode theorientation of four-way switch 78 is changed, thus interchanging therefrigerant connections to the compressor 76, thereby reversing the flowof refrigerant through the system causing the heat exchangers to changefunctions. Referring to waveform t₂ of FIG. 10, drawn from left toright, pressure reduced refrigerant flowing into outside heat exchanger80 (the evaporator) absorbs heat energy from the outside air which isblown through the evaporator 80. The refrigerant flowing through theevaporator remains at its saturation temperature for a majority of thetraverse. As the refrigerant nears the end of the evaporator 80 theaccumulated heat energy that has been absorbed causes the refrigerant totransition to a vapor state. Any further heat energy that is absorbed inthe refrigerant causes the refrigerant temperature to increase beyondthe saturation temperature into a superheated temperature range. Thesuperheated refrigerant flows to the compressor 76 which compresses itto a high pressure, high temperature vapor which is directed to theinside heat exchanger (the condenser) 88. As the high temperature vaporflows into the condenser 88, the temperature of the condenser 88 rapidlyrises to an equivalent temperature. Moisture that had accumulated on theinside coil 88 during the air conditioning mode begins to boil off asthe condenser 88 increases in temperature. The moisture is absorbed byair flowing through condenser 88 into the passenger compartment. Foggingthen occurs when the moisture laden air strikes the cold inside surfaceof the passenger compartment windows.

FIGS. 5 and 11 illustrate an exemplary anti-fogging system forcontrolling the operation of a reversible HVAC system 50 forautomobiles. FIG. 5 as explained earlier in this specificationillustrates a control system for an automotive HVAC system. Using thesame hardware configuration, controller 130 minimizes the effects offogging by gradually increasing the compressor speed at a predeterminedrate and regulating the flow management center operation to ensureefficient use of the evaporator. Although a flow management center 82 isemployed in the preferred embodiment it is within the scope of theinvention to use a pressure reducing device with a separatereceiver/drier. Additionally, the invention encompasses any variablespeed or capacity compressor, even though the compressor in thepreferred embodiment is an electric compressor.

Processor 140 is programmed to control the compressor speed and flowmanagement center operation as depicted in the flowchart of FIG. 11.FIG. 11 provides a general overview of the main system operating modesand the detailed program steps related to the anti-fogging routine. Inthe preferred embodiment of the invention the steps that are included inthe anti-fogging routine 201 are spread throughout a number of programmodules such as the operating mode selection 152, compressor speedcontrol 160, and EXV control 162 (see FIG. 6). Calculated changes to theoutputs that control the speed of compressor 76 and the regulation ofpressure reducing device 116 only occur within the designated modules.To clarify the included steps, they have been brought together andlisted in anti-fog routine 201.

At step 200 the program enters air conditioning mode in which coolingair is blown into the passenger compartment. During air conditioningmode, as a byproduct of the refrigeration process moisture accumulateson the external surface of inside heat exchanger 88. At step 202 ananti-fog flag is set to provide an indication that there is moisture onthe surface of the inside heat exchanger 88. The anti-fog flag willremain set until heat pump mode is entered at step 204. At step 206 theprogram continues into the anti-fog sequence 208 if the anti-fog flag isset, otherwise it branches off to steady-state heat pump mode at step210.

The anti-fog sequence begins with selecting a post-inside heat exchangerair target temperature and a duration of operation at step 212 from atable of values that are represented in the graph. The actualpost-inside heat exchanger air temperature is measured by probe 133. Thetarget temperature is set equal to the ambient plus an offset that isincreased over time. Limiting the post-inside heat exchanger targettemperature to a specified offset above ambient indirectly limits thetemperature of the compressed refrigerant vapor that flows into thecondenser 88. The evaporation rate of moisture located on the insideheat exchanger 88 is directly related to the refrigerant temperature atthe inlet to condenser 88. Therefore, gradually increasing the targettemperature causes a gradual increase in the compressor speed, whichcauses a gradual increase in the compressor discharge pressure, whichresults in a gradual increase in the refrigerant temperature at theinlet to the condenser, thereby limiting the evaporation rate ofmoisture on the condenser 88.

At step 214 the compressor target suction pressure is set to 45 psi.Starting the suction pressure at 45 psi ensures that the startingdischarge pressure and temperature are low enough to preventuncontrolled moisture evaporation from the condenser 88. The suctionpressure is related directly to the speed of compressor 76.

At step 216 a PID controller calculates the new compressor speed settingbased upon the target temperature and previous suction pressure. Thechange in suction pressure from the previous setting is limited toprevent undesirable changes in compressor speed which could lead to highdischarge temperatures and uncontrolled condenser moisture evaporation.Although the preferred embodiment of the invention controls thecompressor speed to regulate the moisture evaporation rate, it is withinthe scope of the invention to control other system parameters such assuction pressure, discharge pressure, or condenser inlet temperature.

If the post-inside heat exchanger target temperature is less than thetarget temperature that correlates to the temperature lever 74 position,then the PTC heater 62 is turned on and the blend door 61 is set to aposition that will enable the HVAC to achieve the temperature levertarget temperature. The required door 61 position is obtained from alookup table that correlates blend door position to differentialtemperature and airflow.

At step 218 the recirculation door 60 is set to the full fresh airposition. Setting the recirculation door 60 to the full fresh airposition in combination with slowly evaporating moisture from thecondenser prevents fogging in the passenger compartment. As moisture isslowly evaporated off of the condenser it is absorbed by the fresh airflowing past the recirculation door 60, through inside heat exchanger88, and into the passenger compartment. The moisture laden air flowinginto the passenger compartment from the outside causes the internal airpressure to increase, acting to drive air out of the compartment throughvents and other unsealed openings. Pushing air out the vents prevents anexcessive amount of moisture laden air from accumulating in thepassenger compartment as well as ensuring that the driest possible airis passed over the inside heat exchanger 88.

The anti-fog sequence continues until controller 130 has executed thetable of values depicted graphically at step 212. Having completed thepredetermined routine, all of the moisture that existed on inside heatexchanger 88 has evaporated and therefore the temperature of therefrigerant entering the condenser 88 no longer needs to be controlled.The anti-fog flag is reset and the heat pump system transitions tonormal steady-state heat pump mode in which the speed of the compressor76 is controlled such that a desired duct outlet temperature as selectedwith temperature lever 74 is attained.

From the foregoing it will be understood that the invention provides asystem which controls fogging when changing modes in a reversible HVACsystem. Additionally, through the use of the anti-fogging method therate of initial heating of the passenger compartment is not compromised.Additionally, the invention permits a system which controls fogging inan HVAC system when initially starting air conditioning mode.

Heating System in a Reversible Air Conditioning and Heat Pump HVACSystem for Electric Vehicles

FIGS. 5 and 12 illustrate an exemplary temperature control system for areversible air conditioning and heat pump HVAC system for an electricautomobile. FIG. 5 illustrates the interconnection of controller 130 toan automotive air conditioning circuit 50. Controller 130 controls thecompressor speed, flow management center 82 operation, and blend door 61positioning based upon inputs from front panel 55, duct 56, and therefrigerant system. The controller 130 is preferably amicroprocessor-based circuit, that includes processor 140 for executinga program, its associated memory 142, an A/D 144 for converting analogsignals into digital inputs, and a driver circuit 146 for interfacingwith system components.

Processor 140 is programmed to control the heating mode selection thatis depicted in the flowcharts of FIGS. 12A and 12B. The heating modeselection programs control the operation of the HVAC circuit 50 during aheating operation. In the preferred embodiment of the invention thesteps that are included in the heating mode selection modules are spreadthroughout a number of program modules such as the operating modeselection 152, compressor speed control 160, and EXV control 162 (seeFIG. 6). Calculated changes to the outputs that control the speed ofcompressor 76 and the regulation of pressure reducing device 116 onlyoccur within the designated modules. To clarify the included steps, theyhave been brought together and listed in the two heating mode selectionmodules.

Heat to the passenger compartment is provided by a combination of theHVAC in heat pump mode and PTC heaters 62 depending on the ambienttemperature and the requested target temperature as selected by theposition of the temperature lever 74.

For ambient temperatures less than 40° F. heat is supplied only by thePTC heater as the reversible HVAC refrigerant system is disabled toprevent icing of the heat exchangers 80 and 88 which would result inreduced airflow and odors in the passenger compartment. At ambienttemperatures greater than or equal to 40° F. heat is supplied by eitherthe heat pump, the PTC heater 62, or the heat pump supplemented by thePTC heater 62.

Referring to FIG. 12A, at step 270 a target temperature is calculatedbased upon the position of temperature lever 74. A lookup table containsvalues that correlate temperature lever position to the targettemperature of the air flowing from the duct outlets 64-68. The creationof a lookup table containing such values is well known in the art. Atstep 272 the target temperature is then compared to the temperature ofair flowing into inside heat exchanger 88. The pre-indoor heat exchangerair temperature is measured by probe 132. If the air temperature atprobe 132 exceeds the target temperature the PTC heater 62 is turnedoff, the heat pump is turned off, and the blend door 61 is set to themax cool position. In the max cool position air bypasses the PTC heaterand flows directly to the duct outlets. During this mode of operationthe outside air which flows into the duct 56 is warmer than thepassenger has requested via the temperature lever 74. To cool theincoming air to the desired temperature the passenger has the option ofenabling air conditioning mode.

For incoming air that is colder than the target temperature thecompressor speed is adjusted by a PID controller at step 276 to drivethe temperature of post inside heat exchanger air to the targettemperature. As compressor speed is increased the refrigerant suctionpressure and temperature decreases enabling the refrigerant to absorb agreater amount of heat from the external air as the refrigeranttraverses the outside heat exchanger (evaporator) 80. The refrigerant isadditionally compressed by the compressor to a greater dischargetemperature and pressure prior to being routed to the inside heatexchanger (condenser) 88. The increased heat load of the refrigerant,obtained from the outside heat exchanger 80, is then transferred to theair flowing through the inside heat exchanger 88. The increased heattransfer causes a commensurate increase in the post inside heatexchanger air temperature, assuming the ambient temperature and air flowrate remains constant.

At step 278 the post inside heat exchanger air temperature is measuredby probe 133 and compared to the target temperature. The post insideheat exchanger air temperature represents the air temperature prior tothe PTC heater. If the air temperature is greater than the targettemperature, then supplemental heat is not required to achieve thetarget temperature. Therefore, at step 280 the controller turns PTCheater 62 off, sets the blend door 61 to the max cool position, andreturns to step 270 to begin another iteration. This is the normaloperating loop during heat mode operation as the controller 130regulates the air temperature to the selected target temperature. Thepost inside heat exchanger air temperature will exhibit normal closedloop operation by fluctuating slightly about the target temperature.

If the measured post inside heat exchanger air temperature is less thanthe target temperature, then the electric heater, PTC heater 62, isturned on. As the air flow rate across the PTC heater 62 increases, theheat output of the device increases thereby transferring a greateramount of heat to the passenger compartment. To regulate the quantity ofheat that is transferred to the passenger compartment blend door 61provides a path for a portion of the air to bypass the PTC heater 62 andrecombine downstream with air that has flowed through the PTC heater 62.By reducing the quantity of air that flows over the PTC heater 62, lessheat is transferred to the air, thereby reducing the commensurateincrease in the temperature of the air, and providing a simple means ofregulating the temperature of the recombined air.

At step 282 the required blend door position to achieve the targettemperature is calculated in a manner known in the art. The requiredeffectiveness represents the amount of PTC heating that is required toraise the temperature of the post inside heat exchanger air to thetarget temperature at the existing airflow across the PTC. At step 284the controller 130 sets the position of blend door 61 and the loopreturns to step 270 to start another iteration. This is the normaloperating loop when supplemental heat from the PTC heater 62 is requiredto raise the duct outlet air to the requested temperature. Each timethrough steps 270, 272, 276, 278, 282, and 284 the position of the blenddoor 61 is varied slightly as the controller 130 responds to changingconditions.

Alternatively, the heating mode selection program can be implemented asillustrated in FIG. 12B. The program illustrated in FIG. 12B isparticularly suitable for operating modes where the overhead energy thatis expended turning on the heat pump or PTC heater 62 exceeds the energyrequired to raise the passenger compartment temperature to the desiredtemperature. At step 300 a forty second timer is started. The timer setsthe time period during which the heat pump attempts to attain the targettemperature. At step 302 the heat pump target temperature is calculatedbased on the position of temperature lever 74. The compressor speed PIDcontroller is adjusted at step 304 to drive the compressor speed towardsattaining the target temperature. At step 306 the heat pump gain iscalculated. The heat pump gain represents the work the heat pumpcontributes to raise the temperature of the passenger compartment underthe existing operating conditions. The heat pump gain is set equal tothe outlet temperature, probe 133, minus the inlet temperature, probe132, divided by the outlet temperature. At step 308 the post-inside heatexchanger air temperature as measured by probe 133 is compared to thetarget temperature calculated at step 302 to determine if the heat pumpis capable of attaining the target temperature. If the heat pump doesattain the target temperature the forty second timer is reset at step310 and the program returns to step 302. Additionally, if the heat pumphas not attained the target temperature but the 40 second timer has nottimed out, the program returns to step 302 to continue to attempt toattain the target temperature. However, if the heat pump does not attainthe target temperature within 40 seconds then at step 314 the measuredvalues for heat pump gain and ambient temperature are stored for lateruse. Although, in the preferred embodiment the heat pump is allowed 40seconds to attain the target temperature, it is within the scope of theinvention that the allowed time may range from about 0 seconds to beyond40 seconds. For example, the heat pump heating capability may becharacterized by factory test or simulation and a number representativeof the capability may be stored in memory for later recall to determineif the heat pump is capable of attaining a target temperature.

At step 316 the heating mode transitions from the heat pump to PTCheater 62 by gradually decreasing the heat pump output and increasingthe PTC heater 62 over a 40 second period. Making a gradual transitionenhances passenger comfort by reducing the noticeability of the changein system operation. At step 318 the stored value for heat pump gain isadjusted for changes in ambient temperature. At step 320 the revisedvalue for heat pump gain is compared to the system gain that representsthe amount of work required to heat the passenger compartment to thetarget temperature. If the system gain exceeds the heat pump gain, thereis insufficient capacity in heat pump mode for heating the passengercompartment, therefore the program remains in PTC heat mode and returnsto step 318. If the heat pump gain exceeds the system gain, the heatpump is capable of supplying the required heat necessary to attain thetarget temperature. The program advances to step 322 and transitionsfrom the PTC heater 62 to heat pump over a 40 second time period,finally returning to heat pump mode at step 300.

From the foregoing it will be understood that the invention provides asystem which minimizes energy consumption during a heating operation ofan automotive HVAC system. Additionally, the method can be employed todynamically update the heating mode selection as operating conditionschange. Also, through the use of the method the energy efficiency of anelectric vehicle is increased. Additionally, the invention provides anenergy efficient method for controlling the passenger compartmenttemperature of an electric vehicle.

Air Handling for HVAC System for Electric Vehicles

Referring to FIGS. 5 and 13, an air handling system for an electricvehicle HVAC system is illustrated. FIG. 5 illustrates theinterconnection of controller 130 to an automotive air conditioningcircuit 50. Controller 130 controls the compressor speed, flowmanagement center 82 operation, and recirculation door 60 positioningbased upon inputs from front panel 55, duct 56, and the refrigerantsystem. Recirculation door 60 may be set to any value from full freshair, through part fresh air with part recirculated air, to fullrecirculated air.

The recirculation door control program 251 is illustrated in FIG. 13.Although FIG. 13 depicts all of the recirculation door programcomponents existing in a single separate program module, it is withinthe scope of the invention for the different elements to be spreadthroughout the system program. In the preferred embodiment of theinvention the steps that are included in the heating mode selectionmodule are spread throughout a number of program modules such as theoperating mode selection 152 and recirc door positioning 154 modules(see FIG. 6). To clarify the included steps, they have been broughttogether and listed in the blend door control module.

When the system is turned-on, step 250 is executed and the recirculationdoor 60 is set to the recirculation position. By starting in therecirculation position less energy is consumed controlling thetemperature of the passenger compartment. In recirculation mode, airfrom within the passenger compartment is routed through the inside heatexchanger 88 before being directed back into the passenger compartment.Therefore to raise the duct outlet air to the desired temperature theheat transferred from inside heat exchanger 88 only has to supplementthe difference between the desired temperature and the temperature ofthe passenger compartment. In fresh air mode, to raise the duct outletair to the desired temperature the heat transferred from inside heatexchanger 88 supplements the difference between the desired temperatureand the temperature of the external air which is flowing into thepassenger compartment.

Having set the recirculation door 60 to its initial position the programcontinues on to step 252 in which the inputs from the front panel 55 areinterrogated to determine if a particular positioning of therecirculation door has been requested. If a recirculation door positionchange has been requested, then at step 254 the recirculation door isset to the requested position at step 254.

In step 256 the program optionally begins an anti-fog sequence. As isexplained above, fogging of the passenger compartment windows may occurwhen the reversible HVAC system 50 switches from cooling mode to heatingmode. During the cooling mode cycle moisture accumulates on the externalsurface of the inside heat exchanger 88 which functions as anevaporator. When the HVAC switches from cooling mode to heating mode therefrigerant flowing into the inside heat exchanger 88, which functionsas a condenser, rapidly increases in temperature. As the refrigerantbegins to raise the temperature of the condenser 88, moisture that hadaccumulated on the inside heat exchanger 88 during the cooling modebegins to boil off. The evaporating moisture is absorbed by air flowingthrough condenser 88 into the passenger compartment. Fogging then occurswhen the moisture laden air strikes the colder windows of the passengercompartment.

At step 258 the air handling procedure during an anti-fog sequence isperformed. The front panel selection for the recirculation door 60position is overridden as the door 60 is set to the full fresh airposition. With fresh air flowing into the passenger compartment the airpressure within the compartment increases, forcing air out of vents anddoor seal cracks. As new fresh air carrying its load of moisture isblown into the passenger compartment, pre-existing moisture laden air isforced out through the vents to the outside environment. Therecirculation door 60 remains in the fresh air position until theanti-fog sequence is completed, at which time the recirculation door isreset to its former position.

In step 260 the program begins an air blow-by sequence. When the vehiclespeed exceeds a predetermined value, such as approximately 42 mph, thepressure from air flowing into the fresh air duct 59 flows not onlythrough the blower 58, but also back up through the recirculation airduct 57. The air flowing back into the recirculation air duct 57bypasses the inside heat exchanger 88 and PTC heater 62 which aredownstream from the recirculation door 60. Therefore, the air flowingback into the recirculation duct is unconditioned external air. Theexternal air could vary from extremely cold dry air during winter monthsto very hot humid air during the summer months. The external air flowsout of the duct inlets and directly onto the passengers in the passengercompartment.

At step 262 the program sets the recirculation door 60 to prevent anundesirable air blow-by event from occurring. The previous setting ofthe recirculation door 60 is overridden and the door is set to the fullfresh air setting. The recirculation air duct 57 is blocked when therecirculation door 60 is in the full fresh air position, therefore thefresh air is forced through blower 58, inside heat exchanger 88, and PTCheater 62. The fresh air is properly conditioned to the desiredtemperature before being blown into the passenger compartment and no airflows back through the recirculation duct 57. Although in the preferredembodiment the recirculation door is set to the full fresh air settingit could alternately be set to the full recirculation air setting, inwhich case the fresh air duct 59 is blocked, preventing fresh air fromflowing into the duct 56. Additionally, although in the preferredembodiment the setting of the recirculation door 60 is independent ofthe prior position of the recirculation door 60, the selection of thefull fresh air setting versus the full recirculation setting could bebased on the position of the recirculation door 60 prior to entering theair blow-by sequence.

From the foregoing it will be understood that the invention provides asystem for selectively overriding the passenger air mixture selectionunder predetermined vehicle operating conditions to permit HVACoperating modes that enhance passenger comfort. Additionally, the systemcan be employed to improve vehicle performance by automaticallyadjusting the air mix during predetermined vehicle operating modes.

System for Cooling Electric Vehicle Batteries

Referring to FIG. 14, a schematic of an automotive HVAC circuit 220 foran electric vehicle is illustrated. The HVAC circuit 220 is analternative embodiment of the invention wherein heat from the batterypack 224 is used to supplement heating of the passenger compartment. Thecircuit 220 is similar to the HVAC circuit 50 illustrated in FIG. 1 withthe addition of a heat exchanger circuit 222 for cooling a battery pack224. A heat exchanger circuit 222 communicates with auxiliary heatexchanger 92 to cool battery pack 224 and controller 130, and includes aheat exchanger 228, a battery pack 224, a reservoir 230, and a pump 232.

Liquid high pressure refrigerant from flow management center 82 flowsthrough expansion valve 226 and shut-off valve 86 into heat exchanger228. Although high pressure refrigerant in the preferred embodiment isobtained from flow management center 82, it is within the scope of theinvention to obtain high pressure refrigerant from other means such as avalve, a receiver/drier, or a reservoir. Additionally, although athermal expansion valve is employed in the preferred embodiment, theprinciples of the invention may be readily extended to other pressurereducing means such as an electronic expansion valve. Shut-off valve 86is included merely to show a possible method of controlling batterycooling by preventing the flow of refrigerant into heat exchanger 228.The refrigerant outlet of heat exchanger 228 is connected to thecompressor 76 suction line such that the vapor is combined withrefrigerant vapor from other system evaporators prior to flowing intothe compressor 76 inlet.

The coolant outlet of heat exchanger 228 connects to battery pack 224which includes the vehicle energy storage batteries. Heat is generatedin the batteries during energy storage and discharge cycles due toenergy losses from converting chemical energy to electrical energy. Heatfrom the batteries is transferred through the battery pack into thecoolant. The outlet of battery pack 224 connects to reservoir 230 whichconnects to the inlet of pump 232. The pump 232 propels the coolantthrough heat exchanger circuit 222. Coolant from the pump 232 flowsthrough controller 130, cooling the system electronics. The heatgenerated by the controller 130 is additionally transferred into thecoolant. The controller 130 controls the operation of HVAC system 220.The temperature of battery pack 224 is sensed by temperature probe 225which provides an input to the Battery Energy Management System (BEMS)234. The BEMS 234 controls the operation of shut-off valve 86 inresponse to the temperature sensed by probe 225.

FIG. 15 illustrates the operation of HVAC circuit 220. In thisembodiment controller 130 sets four-way valve 78 such that the systemheating mode is operational. High pressure, high temperature refrigerantflows from compressor 76 outlet through four-way valve 78 into insideheat exchanger 88 which functions as a condenser. Liquid refrigerantflows from the outlet of condenser 88 through shut-off valve 84 into abi-directional port of flow management center 82. The refrigerant thensplits with a portion flowing from an outlet of flow management center82 to expansion valve 226, and the remainder of the refrigerant flowingout of the expansion valve 116 of the flow management center 226 tooutside heat exchanger 80. Pressure reduced refrigerant flows throughthe outside heat exchanger 80 which functions as an evaporator absorbingheat energy from the outside air flowing through it.

In operation, the refrigerant that flowed from the outlet of flowmanagement center 82 flows through expansion valve 226 and shut-offvalve 86 before entering heat exchanger 228. The pressure reducedrefrigerant that flows through heat exchanger 228 absorbs heat energyfrom coolant that is routed through heat exchanger circuit 222. Theprocess by which heat energy is transferred from the coolant to therefrigerant in heat exchanger 222 is the same as what occurs in insideheat exchanger 88 the functioning of an evaporator described earlier.The coolant in circuit 222 flows through battery pack 224 absorbing heatfrom the vehicle batteries. The coolant then flows through reservoir 230and pump 232 before absorbing additional heat from controller 130 priorto returning to heat exchanger 228. Hot coolant enters the heatexchanger 228 inlet and transfers its heat energy to the pressurereduced refrigerant flowing through the refrigerant line within the heatexchanger 228. The pressure reduced refrigerant transitions to the vaporstate as it absorbs heat energy from the coolant. The vapor staterefrigerant then flows through the four-way switch 78 before combiningwith vapor state refrigerant from outside heat exchanger 80 prior to theinlet to compressor 76.

Coolant continues to circulate through circuit 222 so long as thetemperature of the battery pack 224 remains above 40° F. When thebattery pack 224 temperature decreases below 40° F. the BEMS 234disables shut-off valve 86 interrupting the flow of refrigerant to theheat exchanger 228. Coolant continues to flow through heat exchangercircuit 222 as the temperature of the battery begins to slowly increase.Once the temperature of the battery pack 224 once again rises above 40°F. the BEMS 234 enables shut-off valve 86, reestablishing the flow ofrefrigerant to the heat exchanger 228 and the transfer of heat from theheat exchanger circuit 224 to the HVAC circuit 220 resumes.

Waveform h₁ of FIG. 15 illustrates the heat cycle of HVAC circuit 220.Refrigerant flowing into evaporator 80 initially carries a heat loaddepicted as plateau 238. As the refrigerant flows through evaporator 80it absorbs heat energy from outside air that is blown through theevaporator 80. Meanwhile, refrigerant flowing through heat exchanger 228also carries a heat load depicted as plateau 238. The refrigerantflowing through heat exchanger 228 absorbs heat energy that istransferred from the battery pack 224 of heat exchanger circuit 222. Theheat load of the refrigerant increases to plateau 242 when the vaporstate refrigerant from heat exchangers 80 and 228 combines prior tocompressor 76. The refrigerant heat load further increases to plateau244 when compressor 76 compresses the vapor state refrigerant to a highpressure, high temperature vapor. The stored refrigerant heat energydecreases to plateau 238 as the refrigerant traverses the inside heatexchanger 88 and the heat energy is transferred to air that is blownthrough into the passenger compartment.

Using waste heat from the battery pack to supplement heat energyabsorbed from the outside air for heating the passenger compartmentprovides a number of advantages. It expands the operating conditionsunder which heat mode operation of the HVAC is possible by increasingthe stored energy in the refrigerant. It improves the efficiency of theoverall vehicle system by reducing the need to rely on electric energyto heat the passenger compartment. Where conventional systems wouldexhaust the battery pack waste heat to the external environment and useelectric energy from the batteries to provide supplemental heat to thepassenger compartment, the invention reduces the need for electricalheating by using the waste heat from the batteries to supplement theheat pump system.

During cooling mode the flow of the refrigerant through the main loop isreversed from heat pump mode. Heat from air passing through the insideheat exchanger (evaporator) 88 is absorbed by the refrigerant. Therefrigerant flowing through local-zone heat exchanger 228 continues toabsorb heat from the heat exchanger circuit 222 (refer to FIG. 14). Therefrigerant from the local-zone heat exchanger 228 combines withrefrigerant from inside heat exchanger 88 prior to compressor 76. Therefrigerant is compressed further adding to the heat load and directedto the outside heat exchanger 80 (condenser). As the refrigeranttraverses the condenser 80 the combined heat load is shed to the outsideair that flows through the condenser 80. The refrigerant then flows tothe flow management center 82 and then through the remainder of thecircuit.

From the foregoing it will be understood that the invention provides asystem for increasing the operating range of an automotive heat pumpsystem. Additionally, the invention provides a system for improving theenergy efficiency of an electric automobile. Further, the inventionprovides a system for efficiently distributing the heat energy of anelectric automobile. Also, a method is presented for cooling the batterypack of an electric vehicle.

Advantages of the Invention

From the foregoing it will be understood that the invention provides aflow management device with bi-directional ports in which refrigerantflowing into either port passes through an expansion valve and exits theother port. Additionally, the invention can integrate the receiver/drierfunction into a flow management device with bi-directional ports toprovide the capability of tapping off refrigerant flow for secondarycooling circuits. Also, the present invention decreases the complexityof automotive HVAC systems by integrating a flow management device intothe system to reduce the number of valves required to implement areversible heating and cooling HVAC system. A further capability of theinvention is to provide a centralized flow management center with tapsfor refrigerant to reduce the complexity of automotive HVAC systems thatimplement multi-zone control.

The invention provides a system for improving the steady-state responsetime of an automotive HVAC system. Additionally, the invention permits areduction in the start-up time of an automotive air conditioning system.Also, the invention provides a system for controlling an HVAC systemthat employs a flow management device. The invention further provides asystem for controlling an HVAC system incorporating a centralized flowmanagement center.

The invention provides a system which controls fogging when changingmodes in a reversible HVAC system. Additionally, through the use of theanti-fogging method the rate of initial heating of the passengercompartment is not compromised. Additionally, the invention permits asystem which controls fogging in an HVAC system when initially startingair conditioning mode.

The invention provides a system which minimizes energy consumptionduring a heating operation of an automotive HVAC system. Additionally,the method can be employed to dynamically update the heating modeselection as operating conditions change. Also, through the use of themethod the energy efficiency of an electric vehicle is increased.Additionally, the invention provides an energy efficient method forcontrolling the passenger compartment temperature of an electricvehicle.

The invention provides a system for selectively overriding the passengerair mixture selection under predetermined vehicle operating conditionsto permit HVAC operating modes that enhance passenger comfort.Additionally, the system can be employed to improve vehicle performanceby automatically adjusting the air mix during predetermined vehicleoperating modes.

The invention provides a system for increasing the operating range of anautomotive heat pump system. Additionally, the invention provides asystem for improving the energy efficiency of an electric automobile.Further, the invention provides a system for efficiently distributingthe heat energy of an electric automobile. Also, a method is presentedfor cooling the battery pack of an electric vehicle.

Although certain preferred embodiments of the invention have been hereindescribed in order to afford an enlightened understanding of theinvention, and to describe its principles, it should be understood thatthe present invention is susceptible to modification, variation,innovation and alteration without departing or deviating from the scope,fair meaning, and basic principles of the subjoined claims.

What is claimed is:
 1. A flow management device for regulatingrefrigerant flow in an automotive HVAC system, the flow managementdevice comprising:a housing defining a first port, a second port, and aflow path extending between said first port and said second port tocommunicate a flow of refrigerant therebetween; a first flow sensitivevalve sealingly disposed in said flow path for preventing the flow ofhigh pressure refrigerant from the second port to the first port; asecond flow sensitive valve sealingly disposed in said flow path forpreventing the flow of high pressure refrigerant from the first port tothe second port; a pressure reducing device disposed in said flow pathfor pressure reducing high pressure refrigerant that flows througheither said first flow sensitive valve or said second flow sensitivevalve; and a pressure sensitive valve for preventing the flow of highpressure refrigerant between the first port and the second port, saidpressure sensitive valve having a first position and a second positionsuch that when the pressure sensitive valve is in the first positionpressure reduced refrigerant is permitted to flow from the pressurereducing device to the second port and when the pressure sensitive valveis in the second position pressure reduced refrigerant is permitted toflow from the pressure reducing device to the first port; wherein duringone operating mode high pressure refrigerant flows into the first port,through the first flow sensitive valve, into the pressure reducingdevice whereby pressure reduced refrigerant is emitted, and the pressurereduced refrigerant flows through the multi-function valve and out ofthe second port; and during a second operating mode high pressurerefrigerant flows into the second port, through the second flowsensitive valve, into the pressure reducing device whereby pressurereduced refrigerant is emitted, and the pressure reduced refrigerantflows through the multi-function valve and out of the first port.
 2. Theflow management device of claim 1 further comprising a receiver foraccumulating liquid refrigerant to ensure a continuous supply ofrefrigerant to the pressure reducing valve.
 3. The flow managementdevice of claim 2 further comprising an outlet in the receiver forproviding liquid refrigerant to secondary heat exchanger circuits. 4.The flow management device of claim 1 further comprising a dessicant forfiltering contaminants from the refrigerant.
 5. The flow managementdevice of claim 1 wherein the high pressure refrigerant flowing into thefirst bi-directional port causes the pressure sensitive valve to changeto a position that permits pressure reduced refrigerant to flow out ofthe second bi-directional port.
 6. A flow management device forregulating refrigerant flow in an automotive HVAC system, the flowmanagement device comprising:a housing defining a first port, a secondport, and a flow path extending between said first port and said secondport to communicate a flow of refrigerant therebetween; a first flowsensitive valve sealingly disposed in said flow path for preventing theflow of high pressure refrigerant from the second port to the firstport; a second flow sensitive valve sealingly disposed in said flow pathfor preventing the flow of high pressure refrigerant from the first portto the second port; a pressure reducing device disposed in said flowpath for pressure reducing high pressure refrigerant that flows througheither said first flow sensitive valve or said second flow sensitivevalve; a receiver for accumulating liquid refrigerant to ensure acontinuous supply of refrigerant to the pressure reducing valve; anoutlet in the receiver for providing liquid refrigerant to secondaryheat exchanger circuits; and a multi-function valve for preventing theflow of high pressure refrigerant between the first port and the secondport and for permitting the flow of pressure reduced refrigerant fromthe pressure reducing device to the first port and the second port;wherein during one operating mode high pressure refrigerant flows intothe first port, through the first flow sensitive valve, into thepressure reducing device whereby pressure reduced refrigerant isemitted, and the pressure reduced refrigerant flows through themulti-function valve and out of the second port; and during a secondoperating mode high pressure refrigerant flows into the second port,through the second flow sensitive valve, into the pressure reducingdevice whereby pressure reduced refrigerant is emitted, and the pressurereduced refrigerant flows through the multi-function valve and out ofthe first port.
 7. The flow management device of claim 6 furthercomprising a dessicant for filtering contaminants from the refrigerant.8. The flow management device of claim 6 wherein the multi-functionvalve comprises a first pressure sensitive valve and a second pressuresensitive valve such that pressure reduced refrigerant does not flowthrough the first and the second pressure sensitive valves atsubstantially the same time.